Controller for hydraulic drive machine

ABSTRACT

This invention aims to improve the operability in a hydraulic drive machine. According to a load Pp on work machine actuators 3 and 4, a differential pressure of a discharge pressure Pp of a hydraulic pump and a load pressure PLS is varied. Further, the differential pressure is varied so that the differential pressure between the discharge pressure Pp of the hydraulic pump 2 and the load pressure PLS decreases as the load Pp on the work machine actuators 3 and 4 increases and a rotational speed εE of an engine decreases. Furthermore, the rotational speed εE of the engine 1, discharge pressure Pp of the hydraulic pump 2, the operation amounts S1 and S2 of the various operation levers are detected, respectively, the absorbing torque of the hydraulic pump 2 is set based on the detected rotational speed εE and a target rotational speed εET, and the differential pressure between the discharge pressure Pp of the hydraulic pump 2 and the load pressure PLS is varied based on the detected values and the torque set value. By varying the differential pressure, it is possible to realize an optimal lever operability suitable for the current working conditions, and thus possible to remarkably improve the work efficiency.

TECHNICAL FIELD

The present invention relates to a controller for a hydraulic drivemachine including construction machines such as hydraulic shovels, andrelates in particular to a controller that makes it possible to vary anamount of change in a drive velocity of a work machine actuator perfixed amount of operation of the operation amount of a flow rateoperation valve according to operating conditions of the hydraulic drivemachine.

BACKGROUND ART

Conventionally, control techniques for varying a differential pressurebetween a load pressure of work machine actuators and a dischargepressure of a hydraulic pump according to an externally designated workmode indicating a type of work in order to obtain an operability foroperation levers that corresponds to the nature of the work of aconstruction machine have been disclosed in, for example, JapanesePatent Application Laid-Open No. 2-76904.

With this publicly disclosed technology, it is possible, when the workmode is varied from "normal work" mode to "micro operation" mode, toperform finer work suitable for the "micro operation" mode by making theaforementioned differential pressure smaller than in the case of the"normal work," and by making the change in the drive velocity of thework machine actuators per fixed amount of operation of the operationlevers smaller than in the case of the "normal work."

Japanese Patent Application Laid-Open No. 2-164941 also discloses thistype of control system, and involves effecting control such that theaforementioned differential pressure is reduced in accordance with areduction in the rotational speed of the engine, thereby increasing theso-called metering region, which decreases as the engine rotationalspeed decreases (or, to put it another way, decreases the dead band thatincreases as the rotational speed decreases), and thus improving theoperability of the operation levers.

In this way, these conventional techniques are control methods that varythe differential pressure depending on the work mode or enginerotational speed, that thereby vary the relationship (hereafter referredto as the "operating characteristics") between the work machine actuatorvelocity and the operation amounts of the operation levers, and thatconsequently improve the operability of the operation levers; theseconventional techniques, however, only involve a one-to-onecorrespondence between the change in the differential pressure and thework mode or engine rotational speed, and do not involve controlling thecorrelation with the actual load on the work machine actuators.

However, hydraulic shovels and the like are generally equipped with adevice that effects equivalent horsepower control or the like so thatthe absorbing torque of a hydraulic pump is matched with the engineoutput torque (see FIG. 7(c)), and that controls the discharge quantityof the hydraulic pump according to the so-called PQ curve of FIG. 7(b)so that the absorbing torque value at this matching point is obtained.Such devices are well-known devices which are constructed with, forexample, a TVC valve as a main component.

However, when control is effected in this way so that the torque is keptat or below a certain value, then when the hydraulic pump dischargepressure Pp, that is, the load on the work machine actuators, is high,the hydraulic pump quantity of discharge Qp is low, as shown by the P2(>P1) of FIG. 7(b). For this reason, as shown in FIG. 7(a), compared towhen the load pressure Pp is, on the contrary, low, at P1, when the loadpressure Pp is high, at P2, the work machine drive velocity v isaffected by the low quantity of discharge Qp, and is kept low, as shownby the break line, and the dead band increases and lever operation isgreatly degraded.

With the conventional techniques, moreover, controlling the differentialpressure according to the work mode does not involve effecting controlaccording to the actual work machine drive state pertaining to the workmachine. Specifically, when there are a plurality of work machines, theoperating characteristics required for each one is different, and it isnot possible to meet these requirements by setting a one-to-onecorrespondence between the differential pressure and the work mode. Anexample of a case in which there are different requirements is one inwhich work machine actuators for excavation should be driven as "normalwork," but since the shape of the land is irregular, work machineactuators for travel should be driven as "micro operation." In thiscase, in the past, the work mode designation had to be varied manuallyevery time there was a switch from excavation work to travel work, andvice versa, which was inconvenient in that it complicated operation andincreased the burden on the operator.

With the foregoing in view, the first object of the present invention isto provide a controller that affords better operability than in the pastby controlling the aforementioned differential pressure according to thedrive states and the like of the individual work machines or accordingto the load on the work machine actuators.

The aforementioned prior art, moreover, only involves varying thedifferential pressure in a one-to-one correspondence with the work modeor engine rotational speed, and does not involve effecting control bytaking into account the effects of pressure oil leaks in the actualhydraulic circuit.

Specifically, an increase in the load on the work machine actuators isaccompanied by an increase in pressure oil leakage in the hydraulic pipelines between the work machine actuators and the operation valves (flowrate control valves), and thus by a substantial decrease in the volumeefficiency of the hydraulic pump. A reduction in engine rotationalspeed, moreover, is accompanied by an increase in the ratio of theleakage flow rate to the pump discharge flow rate, and thus by a markeddecrease in the aforementioned volume efficiency. The actual velocity ofthe work machine actuators is therefore decreased, and the relationshipwith the actual operating characteristics is considerably varied. Thedesired operating characteristics consequently cannot be obtained, andthe operability is degraded.

With the conventional techniques, moreover, decreasing the differentialpressure according to the work mode or the engine rotational speed doesnot involve effecting control according to the actual operatingconditions of the operation levers. For example, when all of a pluralityof operation valves (flow rate control valves) are in a neutral positionand a conventional technique is used directly, a phenomenon called"jumping," in which the work machine actuators move suddenly whenoperation lever operation is started, occurs at high engine rotationalspeeds, and at low engine rotational speeds an increase in dead time ordead band occurs when operation lever operation is started; in eithercase, the operability is degraded.

With the foregoing in view, the second object of the present inventionis to provide a device that does not undergo any operability degradationeven in the event of a pressure oil leak, and that does not undergo anyoperability degradation when the operation levers are operated from aneutral position.

The aforementioned conventional techniques involve nothing more thanvarying the differential pressure in a one-to-one correspondence withthe work mode or engine rotational speed, and are not based on thepremise of matching the engine output torque with the hydraulic pumpabsorbing torque.

Accordingly, when applied to a hydraulic shovel or the like, which hasengine output torque limitations, engine failure or the like occurs whenthe load on the work machines becomes great, and work therefore cannotbe continued, which is inconvenient.

With the foregoing in view, the third object of the present invention isto provide a device that makes it possible to prevent inconveniencessuch as engine failure and that improves operability by controlling theaforementioned differential pressure while matching the engine outputtorque and the hydraulic pump absorbing torque.

The aforementioned conventional techniques, moreover, specificallyinvolve providing a valve for differential pressure control to thehydraulic circuit that controls the hydraulic pump, and controlling thehydraulic pump swash plate swash angle by means of this differentialpressure control valve so that a differential pressure corresponding tothe engine rotational speed or work mode may be obtained. Such swashplate control, however, is not based on the premise that the engineoutput torque matches the hydraulic pump absorbing torque.

Consequently, when applied as is to a hydraulic shovel or the like thathas engine output torque limitations, engine failure or the like occurswhen the load on the work machine becomes considerable, and work cannotbe continued, which is inconvenient.

On the other hand, with a hydraulic shovel or the like, there are casesin which a control valve for controlling the pump absorbing torque thatcontrols the hydraulic pump swash plate swash angle so that thehydraulic pump absorbing torque matches the engine output torque isprovided to the hydraulic circuit that controls the hydraulic pump.However, when the absorbing torque control valve and the differentialpressure control valve are both present, but with no interrelationship,and the hydraulic pump is controlled, there are cases in which,depending on the operating conditions, the torque limitations come intoplay, and the operation lever operability is degraded.

With the foregoing in view, the fourth object of the present inventionis to provide a device that makes it possible to prevent suchinconveniences as engine failure, and to thereby improve operability, bycontrolling the differential pressure with the differential pressurecontrol valve, taking into account the aforementioned absorbing torque,while matching the engine output torque with the hydraulic pumpabsorbing torque by means of the absorbing torque control valve.

SUMMARY OF THE INVENTION

To achieve the first object, therefore, the first invention of thisinvention provides a controller for a hydraulic drive machine, whichincludes a hydraulic pump driven by a motor, a plurality of hydraulicactuators driven by the supply of the discharge pressure oil of thehydraulic pump via a pressure oil supply line, and a plurality of flowrate control valves for controlling, in accordance with operationamounts, the flow rate of the pressure oil supplied to the plurality ofwork machine actuators, which controls the discharge flow rate of thehydraulic pump so that a differential pressure between a dischargepressure of the hydraulic pump and a load pressure of the plurality ofwork machine actuators becomes a set value; and which comprises:

pressure detection means for detecting the discharge pressure of thehydraulic pump or the load pressure of the plurality of work machineactuators; and

means for varying the differential pressure set value so that thedifferential pressure set value decreases as the pressure detected bythe pressure detection means increases.

With the structure of the first invention, it is possible to control theexpansion of the operating characteristics dead band that is broughtabout by an increase in the load on the work machine actuators, andthereby possible to improve the lever operability, by means of the factthat an increase in the pressure detected by the pressure detectionmeans is accompanied by a decrease in the differential pressure setvalue.

Similarly, moreover, to achieve the first object, the second inventionof this invention provides a controller for a hydraulic drive machinethat is similar to that of the first invention, which comprises:

pressure detection means for detecting the load pressure of theplurality of work machine actuators or the discharge pressure of thehydraulic pump;

operation amount detection means for detecting the operation amounts ofthe plurality of flow rate control valves; and

means for varying the differential pressure set value so that thedifferential pressure set value decreases as the pressure detected bythe pressure detection means increases, and for varying the differentialpressure set value so that the differential pressure set value decreasesas the operation amounts detected by the operation amount detectionmeans increase while taking as the minimum value the differentialpressure set value determined according to the pressure detection means.

According to the structure of the second invention, the differentialpressure set value corresponding to the pressure detected by thepressure detection means is taken as the minimum value, and an increasein the operation amounts detected by the operation amount detectionmeans is accompanied by a decrease in the differential pressure setvalue. Specifically, when the differential pressure is varied in aone-to-one correspondence by means of the load pressure alone, loadpressure variation is readily generated in the work machine microvelocity region in which the operation amounts are low, and the changein the drive velocity of the work machine actuator caused by thisvariation is discomforting to the operator. When the operation amount islow, therefore, it is possible to remove the discomfort by ensuring thatthe drive velocity variation is not generated, i.e., the differentialpressure does not vary.

Similarly, to achieve the first object, the third invention of thisinvention provides the similar device as that of the first and secondinventions, which comprises:

pressure detection means for detecting the load pressure of theplurality of work machine actuators or the discharge pressure of thehydraulic pump;

operation amount detection means for detecting the operation amounts ofthe plurality of flow rate control valves;

work machine type detection means for detecting a type of work machineactuator currently being driven from among the plurality of work machineactuators; and

means for varying the differential pressure set value based on the typeof work machine actuator detected by the work machine actuator detectionmeans, on the operation amounts detected by the operation amountdetection means, and on the pressure detected by the pressure detectionmeans.

According to the structure of the third invention, the differentialpressure set value is varied based on the type of work machine actuatordetected by the work machine actuator detection means, on the operationamounts detected by the operation amount detection means, and on thepressure detected by the pressure detection means.

Specifically, required operation characteristics which vary depending onan actual work condition of the work machine actuator can be met byvarying the differential pressure according to the type of workmachines.

To achieve the second object, moreover, the fourth invention of thisinvention provides the similar device, which comprises:

pressure detection means for detecting a load pressure of the pluralityof work machine actuators or the discharge pressure of the hydraulicpump;

rotational speed detection means for detecting the rotational speed ofthe motor; and

means for varying the differential pressure set value so that thedifferential pressure set value increases as the pressure detected bythe pressure detection means increases and the rotational speed detectedby the rotational speed detection means decreases.

According to the structure of the fourth invention, the differentialpressure set value varies in such a way that the differential pressureset value increases as the pressure detected by the pressure detectionmeans increases and the rotational speed detected by the rotationalspeed detection means decreases. Specifically, since the differentialpressure set value varies depending on the factors that affect pressureoil leakage in the hydraulic circuit, such as pressure and motorrotational speed, the operability of the flow rate control valve(operation lever for controlling the flow rate control valve) isimproved.

To achieve the second object, moreover, the fifth invention of thisinvention provides the similar device, which comprises:

neutral position detection means for detecting a fact that operatingpositions of the plurality of flow rate control valves are in neutralpositions;

rotational speed detection means for detecting a rotational speed of themotor; and

means for, when the operating positions of all of the plurality of flowrate control valves have been detected by the neutral position detectionmeans to have been in the neutral position, varying the differentialpressure set value so that the differential pressure set value is madelower than the differential pressure set value when any of the pluralityof flow rate control valves is being operated, and so that thedifferential pressure set value decreases as the rotational speeddetected by the rotational speed detection means increases.

According to the structure of the fifth invention, when the operatingpositions of all of the plurality of flow rate control valves have beendetected by the neutral position detection means to have been in theneutral position, then the differential pressure set value changes sothat it is made lower than the differential pressure set value when anyof the plurality of flow rate control valves is operated, and so thatthe differential pressure set value decreases as the rotational speeddetected by the rotational speed detection means increases.Specifically, since the differential pressure in the neutral position issmaller than the differential pressure in a position other than theneutral, and is set according to the rotational speed of the engine,even if the flow rate control valve operation is started from theneutral position, there is no "jumping" phenomenon when operation isbegun at high speeds, and there is no increase in dead time or dead bandat low speeds, so that operability when operation is begun is improved.

To achieve the third object, the sixth invention of this inventionprovides the similar device, in which the operation amounts of theplurality of flow rate control valves, the load pressure of theplurality of work machine actuators or the discharge output of thehydraulic pump, and the rotational speed of the motor are each detected,the absorbing torque of the hydraulic pump is set based on the targetrotational speed of the motor, and the differential pressure set valueis varied in accordance with these detected and set values.

Specifically, according to the structure of the sixth invention, therotational speed of the motor, the hydraulic pump discharge pressure orthe load pressure of the plurality of work machine actuators, and theoperation amounts of the plurality of flow rate control valves are eachdetected, the hydraulic pump absorbing torque is set based on the targetrotational speed of the motor, and the differential pressure set valueis varied in accordance with these detected and set values. Since thedifferential pressure is varied in this way taking the hydraulic pumpabsorbing torque into consideration, a state in which operation cannotbe continued due to inconveniences such as engine failure is prevented.

To achieve the fourth object, the seventh invention of this inventionprovides the similar device, which comprises:

rotational speed detection means for detecting the rotational speed ofthe motor;

discharge pressure detection means for detecting the discharge pressureof the hydraulic pump;

load pressure detection means for detecting the load pressure of theplurality of work machine actuators;

operation amount detection means for detecting the operation amounts ofthe plurality of flow rate control valves;

a controller for setting the hydraulic pump absorbing torque based onthe target rotational speed of the motor, for setting the differentialpressure based on the operation amount detected values of the operationamount detection means, the pressure detected values of the loadpressure detection means or the discharge pressure detection means, therotational speed detected value of the rotational speed detection means,and the set absorbing torque, and for outputting control signalscorresponding to these absorbing torque and differential pressure setvalues;

a torque control valve for controlling a swash plate swash angle of thehydraulic pump, so that the absorbing torque set value is obtained,based on the input from the controller of detected signals correspondingto the discharge pressure detected value of the discharge pressuredetection means and of control signals corresponding to the absorbingtorque set value; and

a differential pressure control valve for controlling the swash plateswash angle of the hydraulic pump, so that the differential pressure setvalue is obtained, based on the input from the controller of detectedsignals corresponding to the pressure detected values of the loadpressure detection means and the discharge pressure detection means andof a control signal corresponding to the differential pressure setvalue.

Specifically, according to the structure of the seventh invention, acontrol signal corresponding to the absorbing torque set value is inputfrom the controller, a detected signal corresponding to the dischargepressure detected value of the discharge pressure detection means isinput, and, based on these signals, the hydraulic pump swash plate swashangle is controlled by the torque control valve so that the absorbingtorque set value is obtained. Meanwhile, a control signal correspondingto the differential pressure set value is input, detection signalscorresponding to the pressure detected values of the load pressuredetection means and the discharge pressure detection means are input,and, based on these signals, the hydraulic pump swash plate swash angleis controlled by the differential pressure control valve so that thedifferential pressure set value is obtained. In this way, the swashplate is controlled by the torque control valve so that the setabsorbing torque is obtained, and the swash plate is controlled by thedifferential pressure control valve, taking the absorbing torque intoconsideration, so that the differential pressure is varied; thisprevents a condition in which operation cannot be continued due toinconveniences such as engine failure, and improves the operability ofthe operation levers.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a circuit diagram which depicts the structure of the workmachine hydraulic circuit in the embodiment of the controller for ahydraulic drive machine which pertains to the present invention.

FIG. 2 is a graph which depicts the relationship between operationamounts of operation levers and drive velocities of work machineactuators in the embodiment.

FIG. 3 is a graph which depicts the relationship between the operationamounts of the operation levers and the drive velocities of the workmachine actuators in another embodiment.

FIG. 4 is a three-dimensional diagram which depicts the contents storedin the controller of FIG. 1.

FIGS. 5(a) to 5(c) are two-dimensional diagrams depicting the contentsof FIG. 4, and are diagrams that are used for explaining the manner inwhich the relationship among the operation amounts of the levers, thepump discharge pressure, and the differential pressure varies dependingon the drive states of the work machines.

FIGS. 6(a) to 6(c) are diagrams which depict the relationshipscorresponding to FIGS. 5(a) to 5(c), respectively, when the hydraulicpump absorbing torque is low.

FIGS. 7(a) to 7(c) are graphs for explaining conventional techniquesrelating to differential pressure control.

FIGS. 8(a) to 8(c) are graphs depicting the manner in which thedifferential pressure varies depending on the engine rotational speedand pump discharge pressure.

FIGS. 9(a) to 9(c) are graphs depicting the manner in which theoperation lever operation amounts, the work machine actuator drivevelocities, and the differential pressure, respectively, vary over time,and are graphs used for explaining the effects of the embodiment in acomparison with conventional techniques.

FIG. 10 is a graph depicting the relationship between the enginerotational speed and differential pressure set value when the operationlevers are in a neutral position.

FIGS. 11(a) to 11(c) are graphs used for explaining an embodiment thatcorrects the pump discharge pressure according to the sum total of theaperture areas of the control valves.

FIG. 12 is a circuit diagram depicting another example of the structureof a work machine hydraulic circuit.

BEST MODE FOR CARRYING OUT THE INVENTION

Next, an embodiment of the controller for a hydraulic drive machinewhich pertains to the present invention is described with reference tothe figures. In the embodiment, furthermore, a hydraulic shovel isassumed to be the hydraulic drive machine.

FIG. 1 shows the structure of the work machine hydraulic circuit thatdrives the various work machines (booms, arms, etc.) of the hydraulicshovel. With the embodiment, moreover, in order to avoid complicatedfigures, only two operation valves corresponding to two types of workmachines are shown.

As shown in the same figure, a variable displacement hydraulic pump 2 isdriven by an engine 1, and the swash angle of a swash plate 2a is variedaccording to the movement of a piston 12a of a swash plate driveregulator 12. The discharge amount D per hydraulic pump 2 revolution(cc/rev) is varied according to the change in the swash angle of thisswash plate 2a. The engine 1 is provided with a rotation sensor 32 fordetecting the rotational speed (rpm) εE of the engine 1, and thedetected signal εE of this rotation sensor 32 is sent to a controller33.

The discharge pressure oil of the hydraulic pump 2 is supplied to a pipeline 9 and to operation valves 7 and 8 via pipe lines 9a and 9b,respectively, formed by the branching of the pipe line 9; Spools of theoperation valves 7 and 8 are driven according to the operation amountsS1 and S2 of operation levers which are not shown, aperture areas A1 andA2 of the operation valves vary depending on the extent of movement ofthe spools, and a flow rate of pressure oil corresponding to this changeis supplied to hydraulic cylinders 3 and 4, respectively, thatconstitute work machine actuators. At this time, the pressure oil thatis circulated from the operation valve 7 is supplied via pipe lines 3aand 3b into the cylinder chamber on the extended side and the cylinderchamber on the retracted side, respectively, of the hydraulic cylinder3, and the hydraulic cylinder 3 is thereby extended and retracted,respectively.

In the same manner, the pressure oil circulated from the operation valve8 is supplied to the cylinder chamber on the extended side and to thecylinder chamber on the retracted side of the hydraulic cylinder 4 vialines 4a and 4b, and the hydraulic cylinder 4 is thereby extended andretracted.

The operation valves 7 and 8 have positions N, M, and L; the pump portto which the pressure oil discharged from the pump 2 is flows in is in aclosed state in the neutral position N, and the pressure oil that flowsthrough the operation valves from the switching position N to theswitching position L or M is throttled by a rotring variable throttle 20that is provided to the spool. In the switching positions L and M,moreover, the throttle 20 has a fixed area, and the load pressure of thehydraulic cylinders 3 and 4, i.e., the pressure at the output side ofpressure reducing valves 25a, 25b, 26a, and 26b that are provided to,respectively, the pipe lines 3a, 3b, 4a, and 4b, is applied to checkvalves 21 and 22 via a port R.

The check valve 21 is connected to a pilot pipe line 23a, and this pilotline 23a is connected to a pilot line 23b. The check valve 22 isconnected to the pilot pipe line 23b. The pilot line 23b is connected toa pilot pipe line 24. The pressure oil on the high pressure PLS side ofthe hydraulic cylinders 3 and 4 is supplied to the pilot pipe line 24via one of the check valves 21 and 22. The pilot pipe line 24 isconnected to the spring position side of the pressure reducing valves25a, 25b, 26a, and 26b, and the load pressure PLS on the high pressureside of the hydraulic cylinders 3 and 4 is therefore applied to thespring position side of the pressure reducing valves 25a, 25b, 26a, and26b. The pressure oil on the input side of the pressure reducing valves,i.e., the pressure on the output side of the operation valves 7 and 8,is applied to a side opposed to the spring side as a pilot pressure. Apipe line 10, moreover, is provided for releasing the pressure oil ofthe operation valves 7 and 8 into a tank 11.

A fixed capacitive hydraulic pump 34 discharges pressure oil at aspecific pressure; this discharged pressure oil is supplied as controlpressure Pc pressure oil to a pilot port 37a of a control valve 37 via apipe line 35 and a control valve 36 (the so-called LS-EPC valve).

Here, the position of the control valve 36 is varied in accordance witha control signal received by an electromagnetic solenoid 36a from acontroller 33, whereby the flow rate of the pressure oil supplied to thepilot port 36a is varied.

A relief valve 38 is furthermore provided to the pipe line 35, andrelief is effected by means of the relief valve 38 when the pressure ofthe discharge pressure oil of the hydraulic pump 34 reaches or exceedsthe pressure set by the relief valve 38.

The pipe line 9 on the discharge side of the hydraulic pump 2 branchesinto a pilot pipe line 14; this pilot pipe line 14 is connected to acylinder chamber on the small diameter side of a regulator 12, and isconnected to a pilot port 37b of the control valve 37. The pilot pipeline 23b is extended, and is connected to a pilot port 37c on the sideon which a spring 37d of the control valve 37 is positioned. For thisreason, the discharge pressure Pp of the hydraulic pump 2 and thecontrol pressure Pc from the control valve 36 are applied to an end ofthe control valve 37 where the spring 37d is not located, the pressurePLS on the high pressure sides of the hydraulic cylinders 3 and 4 isapplied as a pilot pressure to the other end of the control valve 37,where the spring 37d is located, and the energizing pressure of thespring 37d is applied as an offsetting pressure. The position of thecontrol valve 37 is switched according to the differential pressure ofthe pressures applied to each end of the control valve 37, the dischargeflow pressure oil corresponding to the switching position is supplied ordischarged to the cylinder on the large diameter side of the regulator12, and the swash angle of the swash plate 2a is controlled.

In this case, the swash angle of the swash plate 2a is controlled sothat the differential pressure ΔPLS between the hydraulic pump pressurePp and the cylinder load pressure PLS is maintained at the set value, asdescribed below. In this case, the differential pressure ΔPLS set valueis varied according to the control pressure Pc, i.e., the control signalsent to the electromagnetic solenoid 36a from the controller 33.

The relationship between the pressures Pp and PLS and the dischargeamount (volume) D of the hydraulic pump 2 at this time is expressed byformula (1) below.

    D=C·A·√(Pp-PLS)                   (1)

Here, C is a constant, and A is the aperture area of the throttle 20.

A fuel injection pump 38 and a governor 39 are both provided to theengine 1. It is driven by a motor 40 or a fuel control lever 39a of thegovernor 39, and the drive position of the lever 39a is detected by aposition sensor 41. The detection signal of the position sensor 41 issent to the controller 33 as a feedback position signal during the drivecontrol of the motor 40.

A throttle dial 42 sets the target rotational speed of the engine 1, anda throttle signal corresponding to the target rotational speed εTH issent to the controller 33. A monitor panel 43, moreover, selects anddesignates the work mode M effected by the hydraulic shovel, i.e., a"heavy excavation" mode M1, an "excavation" mode M2, an "adjustment"mode M3, or a "micro operation" mode M4, and a signal indicating theselected work mode M1, M2, M3, or M4 is sent to the controller 33.

A pump pressure sensor 44 is disposed in the pipe line 14, which detectsthe pressure of the pressure oil within the pipe line 14, i.e., thedischarge pressure oil Pp of the oil pressure pump 2. The detected valuePp is then applied to the controller 33.

Operation amount sensors 45 and 46 for detecting stroke operationamounts (hereafter referred to as operation amounts) S1 and S2 areprovided to the operation valves 7 and 8, respectively, and the detectedvalues S1 and S2 are sent to the controller 33.

The controller 33 outputs a drive control signal to the motor 40 basedon the various signals that have been input, and thereby controls theoutput torque of the engine 1. Specifically, as shown in FIG. 7(c), adrive control signal is sent to the motor 40 so that regulation lines11, 12, 13, and so on corresponding to the input target rotational speedεTH and the current engine rotational speed εE detected by the enginerotation sensor 32 are established, and the fuel control lever 39a isoperated.

Meanwhile, the controller 33 effects processing such as that describedbelow based on the various input signals, and outputs the control signalobtained as a result to the solenoid 36a of the control valve 36 tocontrol the swash angle of the swash plate 2a of the hydraulic pump 2,i.e., the discharge amount D (cc/rew) of the hydraulic pump 2, via thecontrol valve 37 and the regulator 12. In this case, the controller 33outputs a control signal that sets the absorbing horsepower of thehydraulic pump 2 to a fixed value. Specifically, the hydraulic pump 2outputs to the control valve 36 a control signal such that a fixedhorsepower corresponding to the input work mode M1 . . . is obtained,and thereby controls the swash plate 2a of the hydraulic pump 2 via thecontrol valve 37. In this way, the matching point moves to the point ofoptimal efficiency according to the current load conditions (see FIG.7(c)).

Meanwhile, the controller 33 outputs a control signal so that thedifferential pressure ΔPLS set in the manner described below isobtained. Specifically, the controller 33 effects both the control ofthe pump absorbing horsepower and the control of the differentialpressure by means of the same control signal; in this case, the controlpressure Pc applied to the pilot port 37a of the control valve 37 variesaccording to the control signal sent to the solenoid 36a of the controlvalve 36, whereby the differential pressure ΔPLS is varied. With thisembodiment, this differential pressure ΔPLS is varied according tovarious controls, as described below, to improve the operability of theoperation levers (not shown in the figures) of the operation valves 7and 8. The variable control of this differential pressure ΔPLS isdescribed in detail below.

First Control

This first control aims to improve the operability by varying thedifferential pressure ΔPLS depending on the load currently on the workmachine actuator.

Specifically, in the system shown in FIG. 1, the output torque of theengine 1 and the absorbing torque of the hydraulic pump 2 are generallymatched at the matching point, as shown in FIG. 7(c), and the dischargequantity Q (cc/min) of the pump 2 is controlled according to a PQ curvesuch as that shown in FIG. 7(b), so that the absorbing torque at thistime can be obtained. In this way, the engine failure of the engine 1 isprevented by means of equivalent horsepower control. As shown in FIG.7(b), however, an increase in the load PLS on the work machine actuators3 and 4, i.e., in the discharge pressure Pp of the hydraulic pump 2, isaccompanied by a decrease in the pump discharge quantity Q. For thisreason, when the load is considerable, the engine failure preventionfunction acts as a limiter on the pump discharge amount (volume)(cc/rev).

FIG. 7(a) shows the general relationship between the operation amounts S(S1, S2) of the operation levers and the drive velocities v (v1, v2) ofthe work machine actuators 3 and 4; when the work machine actuator witha considerable load is driven and the discharge pressure Pp increasesfrom P1 to P2 (see FIG. 7(b)), this is accompanied by the control of thedischarge amount, as shown by the dotted line, and thus by the controlof the drive velocity, which results in the enlargement of the so-calleddead band (dead stroke).

With this embodiment, therefore, a control signal is output to thecontrol valve 36 so that an increase in the discharge pressure Ppdetected by the pump pressure sensor 44, i.e., in the load PLS of thework machine actuators 3 and 4, is accompanied by a decrease in thedifferential pressure ΔPLS.

Specifically, FIG. 2 depicts the relationship between the work machineactuator drive velocities v1 and v2 and the operation amounts S1 and S2pertaining to the "first control"; because of the fact that an increasein the load Pp is accompanied by a decrease in the differential pressureΔPLS, even when the load Pp is considerable, there is no shift to thecharacteristics (b) with a considerable dead band, but there is a shiftto the characteristics (c) with a low gradient, and the dead band issmall, as is the case with the characteristics (a) where the load Pp issmall; this allows good lever operability to be maintained.

Second Control

With the first control described above, the differential pressure ΔPLSis varied according to the pump discharge pressure Pp; however, when thedifferential pressure ΔPLS is varied in a one-to-one correspondence bymeans of the pump discharge pressure Pp alone, load variations arereadily brought about in the work machine actuator micro velocity regionin which the operation amounts are low, and this load variation resultsin a variation in the velocities v1 and v2, which causes the operatorsome discomfort. With this embodiment, therefore, a control signal isoutput to the control valve 36 so that increases in the operationamounts S1 and S2 detected by the operation amount sensors 45 and 46,respectively, are accompanied by a decrease in the differential pressureΔPLS, with the differential pressure determined by the pump dischargepressure Pp being taken as the minimum value; the aforementioneddiscomfort is thereby eliminated.

The operating characteristic (c)' shown in FIG. 3 depicts therelationship between the work machine actuator drive velocities v1 andv2 and the operation amounts S1 and S2 which pertain to th "secondcontrol"; the larger the operation amounts S1 and S2, the smaller thedifferential pressure ΔPLS, and the drive velocities v1 and v2 are notsuddenly limited with an increase in the operation amounts S1 and S2, asis shown by the characteristic (b), but the drive velocities v1 and v2vary in such a way that they gradually approach the limiting values ofthe drive velocities v1 and v2 as the operation amounts S1 and S2 rise.In this case, when the operation amounts S1 and S2 of the work machineactuators are small, the differential pressure ΔPLS is large, and when,for example, the load is high, the characteristics are virtuallyunchanged from the characteristics (a) for a low load. In short, in themicro velocity region, the differential pressure is virtually unchangedby variations in the load, and it is possible to obtain favorableoperability in which no discomfort is felt due to variations in thedrive velocities v1 and v2. Ultimately, an increase in the operationamounts S1 and S2 is accompanied by a decrease in the differentialpressure ΔPLS (the gradient becomes small due to the characteristic),until finally the differential pressure (minimum value) set by the pumpdischarge pressure is reached, and a dead stroke condition results.

The differential pressure ΔPLS is furthermore varied depending on thelarger of the operation amounts S1 and S2 of the operation valves 7 and8. The operation amounts S1 and S2, moreover, indicate the apertureareas A1 and A2 of the operation valves, and may thus be used to detectthis.

Third Control

With the above-described first and second controls, the differentialpressure ΔPLS is varied according to the pump discharge pressure Pp orthe operation amounts S1 or S2 (whichever is the larger) so as to obtainthe desired operating characteristics in a one-to-one correspondence;with a hydraulic shovel equipped with a plurality of work machines,however, required operating characteristics are different depending onwhich of the work machines is driven. With this third control,therefore, the direction in which a work machine is driven is detectedby work sensors 45 and 46, operating characteristics are selectedaccording to this detected work machine (boom, arm, or the like), andcontrol is effected so as to obtain these selected operatingcharacteristics, whereby the characteristics described above aresatisfied. In this case, each control characteristic is stored inadvance in a memory (not shown in the figure) in the controller 33, asshown in FIG. 4, as a three-dimensional map which depicts therelationship among the differential pressure ΔPLS, the pump dischargepressure Pp, and the lever operation amount Si (or the aperture area Aiof the operation valve) for each type of driven work machine i.Furthermore, the shape of the three-dimensional map E of FIG. 4 inpractice differs for each work machine drive state, and FIG. 4 isnothing more than a simple example.

FIGS. 5(a) to 5(c) are two-dimensional representations of thethree-dimensional map E shown in FIG. 4 that have been separated intocases in which the load Pp is low (FIG. 5(a)), the load Pp is of mediumvalue (FIG. 5(b)), and the load Pp is high (FIG. 5(c)) and that show thedrive states of the work machines, i.e., boom elevation (dotted line C),adjustment mode M3 arm excavation (dot-dash-dot line B), and other cases(solid line A).

As is clear from these figures, when the work machine drive state hasbeen detected as being "boom elevation," the differential pressure ΔPLSis fixed, regardless of the load Pp detected value. This is due to thefact that since the load is extremely high during boom elevation, thereare no problems from the standpoint of operability if the operatingcharacteristics are determined by means of the work machine drive statealone, without regard to the load Pp detected value.

The operating characteristics (differential pressure) that should beselected may also be varied depending on the size of the absorbingtorque of the hydraulic pump.

For example, when it has been made clear by the controller 33 that theabsorbing torque of the hydraulic pump 2 has been set low, the A, B, andC characteristics shown in FIGS. 5(a) to FIG. 5(c) become thecharacteristics A', B', and C' shown in FIGS. 6(a) to 6(c). In thiscase, the reason that characteristic A' in FIG. 6(a) shows a decreasingdifferential pressure ΔPLS as the operation amount Si of the leverincreases is that full lever operation is dependent on the powerrestrictions of the engine 1.

Even in the same work mode, when the work machine actuator forexcavation work, such as the boom, is driven, and a work machineactuator for travel is driven, as work machine actuators, the operatingcharacteristics required by each of the work machine actuators aredifferent, and it is sometimes inconvenient to determine the operatingcharacteristics based on a one-to-one correspondence with the work mode.Specifically, in operation in "excavation mode" M2 as well, as far asthe drive velocity of the actuators for the boom, arm, and the like areconcerned, although the aim is to effect drive at an ordinary drivevelocity corresponding to "excavation mode," in the movements during theexcavation work, because of the irregular shape of the earth's surface,for the sake of safety there is sometimes the contradictory requirementthat the speed difference between ascending and descending slope shouldbe made small, and the drive velocity of the actuator for travel shouldbe made low, i.e., drive should be effected at a drive velocitycorresponding to the "micro operation mode" M4.

In this case, when the excavation mode M2 is selected as the currentwork mode, a determination of which work machine actuator is currentlybeing driven is made based on the detected values of the operationamount sensors 45 and 46; if travel is currently under way, operatingcharacteristics with a low gradient and a low drive velocity(corresponding to the micro operation mode M4) similar to thecharacteristics (c) of FIG. 2 are selected to effect micro operation. Onthe other hand, when it has been determined that the boom or the like iscurrently being driven and that excavation is under way, operatingcharacteristics such as those shown by characteristics (a) of FIG. 2,for which the slope is greater and the drive velocity is higher, andwhich correspond to the currently selected "excavation mode" M2, areselected. When it has been determined that the actuator for excavationand the actuator for travel are being driven simultaneously as workmachine actuators, moreover, for the sake of safety operatingcharacteristics such as those shown in the characteristics (c) of FIG.2, for which the drive velocity becomes low, are selected.

A generalization of the third control goes as follows: a function G thathas as variables the pump discharge pressure Pp and the operationamounts S1 to Sn (1 to n indicate various work machines) is firstdetermined, and this is used to determine the differential pressureΔPLS.

    ΔPLS=G(Pp, S1 to Sn)                                 (2)

In this case, the map E of FIG. 4 shows the function G in athree-dimensional manner.

As shown by formula (3) below, moreover, the differential pressure ΔPLSthat is determined by means of this function G may be bound by therestriction that the maximum differential pressure ΔPLS_(max) set inadvance at a time of low load must not be exceeded.

    ΔPLS=min(G(Pp, S1 to Sn), ΔPLS.sub.max)        (3)

This maximum differential pressure ΔPLS_(max) is the differentialpressure that determines the cycle time; when it is low the cycle timeis delayed. For example, it is possible to make the cycle timecorrespond to the work mode by varying the maximum differential pressureΔPLS_(max) depending on the work mode.

In the embodiment, moreover, although the differential pressure isvaried based on the discharge pressure Pp of the hydraulic pump 2, it isalso possible to vary the differential pressure based on the load of thework machine, and it is of course also possible to vary the differentialpressure based on the load PLS of the work machine.

It is of course also possible to suitably combine the first throughthird controls described above with conventional techniques relating todifferential pressure control (e.g., Japanese Laid-Open PatentApplications 2-76904 and 2-164941). In this case, even when, forexample, the differential pressure is varied based on the load Pp sothat the characteristics (c) shown in FIG. 2 are obtained, it ispossible to consider varying the amount of this change depending onwhich work mode (M1 . . . ) has been selected by the monitor panel 43.

As described above, this embodiment involves varying the differentialpressure ΔPLS according to, for example, the load Pp on the work machineactuator, and thus makes it possible to obtain optimal operabilitysuited to the current working conditions and to dramatically improveworking efficiency over that achieved in the past.

Fourth Control

With this fourth control, the differential pressure ΔPLS is varieddepending on the engine rotational speed and the load currently on thework machine actuator to effect control that does not sacrifice leveroperability even in the case of a pressure oil leak described above.

Generally, the effects of a pressure oil leak on the operatingcharacteristics are said to be proportional to the ratio qL/Q of theleakage quantity qL in the hydraulic pump 2 hydraulic oil pipe line tothe discharge quantity Q (cc/min). When this ratio qL/Q becomes high,the practical volume efficiency of the hydraulic pump 2 is degraded, theactual velocity of the work machine actuator falls, and the operationlever operating characteristics vary from the desired operatingcharacteristics in the direction of decreasing differential pressure.Making the ratio qL/Q low therefore makes it possible to maintain thedesired operating characteristics, and thus makes it possible tocomplete the operation without sacrificing lever operability.

Here, the pump discharge quantity Q is defined by

    Q=D·εE                                    (4)

and is proportional to the engine rotational speed εE. On the otherhand, the leakage quantity qL itself is known to be proportional to theloads on the work machine actuators 7 and 8, i.e., to the dischargepressure Pp of the hydraulic pump 2. Consequently, the ratio qL/Q isexpressed by

    qL/Q=Pp/εE                                         (5)

and since the ratio qL/Q ultimately increases as the hydraulic pumpdischarge pressure Pp increases, in order to prevent a consequentdecrease in differential pressure, correction is made in the directionof increasing the differential pressure as the pressure Pp increases,which makes it possible to maintain the desired operatingcharacteristics; since the ratio qL/Q increases as the engine rotationalspeed εE decreases, moreover, in order to prevent the consequentdecrease in differential pressure, correction is made in the directionof increasing the differential pressure as the rotational speed εEdecreases, which makes it possible to maintain the desired operatingcharacteristics.

FIGS. 8(a) to 8(c) show, for the implementation of this first control,the relationship between the operation amounts S1 and S2 (or theoperation valve aperture areas S1 and S2) of the operation levers andthe differential pressure ΔPLS, separated into cases in which the pumpdischarge pressure Pp is low (FIG. 8(a)), cases in which the pumpdischarge pressure Pp is of medium value (FIG. 8(b)), and cases in whichthe pump discharge pressure Pp is high (FIG. 8(c)), for cases in whichthe engine rotational speed εE is low (dot-dash-dot line A) and cases inwhich the engine rotational speed εE is high (solid line B).

As is clear from FIGS. 8(a) to 8(c), as the pump discharge pressure Ppincreases from FIG. 8(a) to FIG. 8(b), and then to FIG. 8(c), thedifferential pressure ΔPLS increases, and as the engine rotational speedεE decreases from B to A, the differential pressure ΔPLS is set at ahigh level.

The contents of FIGS. 8(a) to 8(c) are stored in advance in a memory,not shown in the figures, in the controller 33, the differentialpressure ΔPLS corresponding to the detected values of FIG. 8(a), FIG.8(b), and FIG. 8(c) is fetched based on the engine rotational speedεEdetected by the rotation sensor 32 and the pump discharge pressure Ppdetected by the pump pressure sensor 44, and a control signal is outputto the control valve 36 so that this differential pressure ΔPLS isobtained. As a result, the lever operating characteristics are notvaried even in the event of a pressure oil leak, and the desiredoperating characteristics are maintained.

With this fourth control, furthermore, although the differentialpressure is varied based on the discharge pressure Pp of the hydraulicpump 2, it is also possible to vary the differential pressure based onthe load on the work machine, and it is of course also possible to varythe differential pressure based on the load PLS of the work machine.

Fifth Control

With this fifth control, when the operation valve is in a neutralposition, the differential pressure ΔPLS is lowered below the setdifferential pressure for positions other than neutral, and is variedaccording to the engine rotational speed; this is effective atpreventing the generation of such problems as "jumping" at high enginerotational speeds and "dead time increase" at low engine rotationalspeed, which have been described above, and thereby improves operabilitywhen lever operation is begun.

As described above, Japanese Patent Application Laid-Open No. 2-164941provides an improvement in operability by effecting control in such away that the differential pressure PLS is decreased with a decrease inengine rotational speed; when all of the operation valves are operatedin a neutral position N, however, and the aforementioned control iseffected in this condition, then the differential pressure ΔPLS becomesconsiderable when operation lever operation is begun, as shown by G inFIG. 9(a), and at high engine rotational speeds, as shown by H in FIG.9(b), and this in turn results in the "jumping" phenomenon, in which thework machine actuator drive velocity increases suddenly. This is causedby the fact that there is no difference between the differentialpressure set in the neutral position N and the differential pressure setwhen control is effected for a position other than the neutral positionN, and is also caused by a sudden increase in the differential pressureΔPLS when lever operation is begun, as shown by I in FIG. 9(c).

The differential pressure when the operation levers, i.e., the operationvalves 7 and 8, are in the neutral position N is ΔPLSn, and thedifferential pressure when the operation levers are in an operated stateother than the neutral position N is ΔPLSa; when the neutral positiondifferential pressure ΔPLSn is lower than the control differentialpressure ΔPLSa, as in

    ΔPLSn>ΔPLSa                                    (6)

then the differential pressure increases along a transitional, gentleslope such as that shown by the dotted line J in FIG. 9(c), and the"jumping" phenomenon is eliminated, as shown by the dotted line K inFIG. 9(b).

At low engine rotational speeds, on the other hand, with theconventional method the differential pressure ΔPLS becomes low both atthe neutral position N and at low engine rotational speeds, there is noincrease in work machine actuator drive velocity when lever operation isbegun, as shown by the dot-dash-dot line L in FIG. 9(b), and an increasein dead time and dead band occurs. Consequently, in the neutral positionN, as opposed to the other positions, the differential pressure ΔPLSn isincreased in accordance with a decrease in the engine rotational speedεE, as shown in FIG. 10, and it is thereby possible to remove suchinconveniences as the increase in dead time. As shown in this FIG. 10,moreover, the differential pressure ΔPLSn is varied so that it decreaseswith an increase in engine rotational speed εE, so that it is possibleto effectively prevent the "jumping" phenomenon that becomes marked withan increase in engine rotational speed.

Ultimately, as shown by Formula (6) and FIG. 10, when either of theoperation valves 7 and 8 is in the neutral position, the differentialpressure ΔPLSn is set so that it is smaller than the differentialpressure ΔPLSa when either of the operation valves 7 and 8 is operated,and so that it decreases as the engine rotational speed εE increases;both of the aforementioned inconveniences are thereby eliminated, andoperability when lever operation is begun can thus be improved.

The contents of Formula (6) and FIG. 10 are stored in advance in amemory, not shown in the figure, in the controller 33, and a check ismade to detect if either of the operation valves 7 and 8 are in theneutral position N, and when this neutral position N is detected, thenthe differential pressure ΔPLSn corresponding to the output εE of therotation sensor 32 is fetched from the memory, and a control signal isoutput to the control valve 36 so that this differential pressure ΔPLSnis obtained. As a result, the "jumping" phenomenon and the like areeliminated when lever operation is begun, and an operability superior tothat achieved in the past is realized.

This fifth control is thus clearly suitable not only for cases in whicha conventional technique for reducing the differential pressureaccording to a decrease in engine rotational speed is used, but also forcases in which the differential pressure is set at the time of leveroperation without consideration of engine rotational speed.

Sixth Control

With this sixth control, the rotational speed εE of the engine i and thedischarge pressure Pp of the hydraulic pump 2, i.e., the load pressurePLS of the work machine actuators 3 and 4 and the operation amounts S1and S2 of the operation valves 7 and 8, are detected, the absorbingtorque τ of the hydraulic pump 2 is set by means of equivalenthorsepower control based on the detected rotational speed εE and thetarget rotational speed εTH of the engine 1, and the differentialpressure ΔPLS is varied in accordance with these detected values and thetorque set value τ, whereby control limiting the absorbing torque of thehydraulic pump 2 is effected, inconveniences such as engine failure areprevented, and good lever operability is obtained.

In general, the relationship shown in Formula (7) below holds among thedifferential pressure ΔPLS, the absorbing torque τ of the hydraulic pump2, the sum total A of the aperture areas of the operation valves 7 and8, the discharge pressure Pp of the hydraulic pump 2, and the rotationalspeed εE of the engine 1.

    √(ΔPLS)=εE·τ/(k·Pp·A) (7)

Here, A=A1 to An (where 1 to n indicate operation valves; A1+A2 in thisembodiment). Formula (7) is obtained as described below. Specifically,the relationship Q=D·εE holds between the capacity D and the dischargequantity Q (cc/min) of the hydraulic pump 2, and the absorbing torque τof the pump 2 is expressed by τ=D·Pp=τ(τE·εTH). Thus, Q=C·A·√(ΔPLS)holds according to Formula (1). Formula (7) is thus obtained byeliminating Q and D from these formulas. Furthermore, since the pumpdischarge pressure Pp and the actuator load pressure PLS are essentiallythe same, PLS can be used in place of Pp in Formula (7).

The maximum value for the discharge quantity Q of the hydraulic pump 2is determined when the operation valves 7 and 8 are operated up to themaximum operation amounts at the maximum rotational speed of theengine 1. The differential pressure ΔPLS obtained through Formula (7) byfirst determining this discharge quantity Q maximum value and by takingthe corresponding maximum differential pressure as ΔPLS_(max) must notexceed the maximum differential pressure ΔPLS_(max). Ultimately, thedifferential pressure ΔPLS is determined by means of Formula (8).

    ΔPLS=min({εE·τ/(k·Pp·A}2, ΔPLS.sub.max)                                       (8)

The εE, Pp, and A of εE·τ/(k·Pp·A) in Formula (8) can be obtained fromthe detected values of the corresponding sensors, and τ is obtained bysetting the absorbing torque τ of the hydraulic pump 2 according toequivalent horsepower control based on the detected rotational speed εEand the target rotational speed εTH of the engine 1. The aperture areasum total A may be obtained as the sum of the outputs S1 and S2 of theoperation amount sensors 45 and 46, or may be obtained as the larger ofthe outputs S1 and S2 of the operation amount sensors 45 and 46.

As described above, because the set horsepower varies (the equivalenthorsepower curve shown in FIG. 7(c) differs) depending on the work mode(M1 . . . ), and the absorbing torque τ set thereby thus also varies, itmay be so arranged that the function εE·τ/(k·Pp·A) on the right side ofFormula (7) is prepared for each work mode (M1 . . . ) as a function inwhich the engine rotational speed εE . . . are variables, the functioncorresponding to the selected work mode (M1 . . . ) is selected, and thedifferential pressure ΔPLS is calculated based on this selectedfunction.

Furthermore, it is also possible to prepare the function εE·τ/(k·Pp·A)for each drive state of the work machine actuators 3 and 4, according towhich work machine is being driven in which direction. Since theabsorbing torque to be set varies depending on the drive state, it isnecessary, for example, in the case of boom elevation, to set theabsorbing torque τ high because the load is high, and in the case ofbucket operation, the absorbing torque may be set low because the loadis comparatively low. Which work machine is being driven in whichdirection can moreover be detected based on the outputs of the operationamount detection sensors 45 and 46.

Since the maximum differential pressure ΔPLS_(max) also varies dependingon the drive state of the work machine actuators 7 and 8 and on theselected work mode (M1 . . . ), it can also be determined based onthese.

With this sixth control, therefore, the aforementioned function isselected based on the selected work mode (M1 . . . ), the type ofcurrently driven actuators 3 and 4 detected by the operation amountsensors 45 and 46, and the drive direction thereof, and the substitutionof the engine rotational speed εE . . . into this selected functionallows the differential pressure ΔPLS of Formula (7) to be determined.Meanwhile, the differential pressure maximum value ΔPLS_(max) isdetermined based on the selected work mode (M1 . . . ), the type ofcurrently driven actuators 3 and 4 detected by the operation amountsensors 45 and 46, and the drive direction thereof, the smallerdifferential pressure ΔPLS is determined by Formula (8), and a controlsignal is output to the control valve 36 so that this determineddifferential pressure ΔPLS is obtained.

The operating characteristics (a) and (c) of FIG. 2 express therelationship between the operation amounts S1 and S2 and the workmachine actuator drive velocities v1 and v2 for cases in which the loadis low and cases in which the load is high, respectively, according tothis sixth control; since the differential pressure ΔPLS decreases asthe load Pp increases according to Formula (7), there is no shift fromthe characteristics (a) to the characteristics (b), in which there is aconsiderable dead band, even when the load Pp is high, but there is ashift towards the characteristics (c), in which there is a low gradient,so that the dead band is kept low, as is the case with characteristics(a), in which the load Pp is low, and good operability is thusmaintained. Moreover, since the equivalent horsepower control of theabsorbing torque of the hydraulic pump 2 is effected at the same time,no inconveniences such as engine failure occur. The characteristics (b)shown by the dotted line in FIG. 2, furthermore, depict cases in whichcontrol is not effected based on Formula (7), from which it is seen thatthe dead band is enlarged and the operability is degraded due to atorque limitation encountered when the load is high.

Seventh Control

With the sixth control, since the differential pressure ΔPLS is variedaccording to Formula (7), good operability that is suitable for the loadcurrently on the work machine can be obtained; with this seventhcontrol, however, the aim is to realize more precise control bycorrecting the load Pp of Formula (7).

FIG. 11(c) shows the relationship between the pump discharge pressure Ppand the pump discharge quantity Q; since the PQ curve is generallyapproached as the sum total A of the aperture areas decreases, i.e., asthe operation amounts S1 and S2 of the operation levers decrease,variations in the actual pressure Pp result, as shown by G, invariations in the quantity Q, and thus in variations in the differentialpressure, which has an adverse effect on operability.

Ultimately, as shown in FIG. 11(a), this seventh control allows the pumpdischarge pressure detected value Pp to be corrected so that thedischarge pressure Pp' gradually increases, as shown by the dot-dash-dotline I and the dotted line H, .as the aperture area sum total Adecreases. In FIG. 11(a), the solid line J depicts the relationshipbetween the detected value Pp and the corrected value Pp' when theaperture area sum total A is at its maximum value A_(max) ; when theaperture area sum total A is at its maximum, there is no degradation inoperability, and the detected value Pp is thus not corrected. When theaperture area sum total A is higher than the minimum value A_(min) andlower than the maximum value A_(max), correction is performed, asindicated by the dotted line H, and when the aperture area sum total Ais at the maximum value A_(max) then, as shown by the dot-dash-dot lineI, the corrected value is made larger than in the case of the dottedline H, resulting in a degradation of operability.

The reason that the corrected value decreases as the detected valueincreases is that, since the variation width of the flow rate Qdecreases as the pump pressure Pp increases, as is clear from FIG.11(c), the variation in the differential pressure decreases, and notthat much correction is required.

The contents of FIG. 11(a) may be expressed, as FIG. 11(b), in which therelationship among the pump pressure detected value Pp, the aperturearea sum total A, and the corrected value Pp' is represented as athree-dimensional map K, and correction may be carried out according tothis three-dimensional map K.

Accordingly, with this seventh control, the contents of FIG. 11(a) orFIG. 11(b) are stored in advance in the memory, not shown in thefigures, in the controller 33. The corresponding corrected value Pp' inFIG. 11(a) or FIG. 11(b) may thus be fetched based on the detected valuePp of the pump pressure sensor 44 and the detected values S1 and S2 ofthe operation amount sensors 45 and 46. In this case, the aperture areasum total A may be determined from the sum total of the operationamounts S1 and S2, or may be determined to be the larger of theoperation amounts S1 and S2.

The thus obtained corrected value Pp' is then used to correct Formula(8) to Formula (9) below.

    ΔPLS=min({εE·τ/(k·Pp'·A)}2, ΔPLS.sub.max)                                       (9)

A control signal for obtaining this corrected differential pressure ΔPLSis then output to the control valve 36. As a result, operability in themicro velocity region of the operation levers is further improved.

Eighth Control

With the sixth control, the differential pressure ΔPLS is variedaccording to Formula (7), and good operability that suits the loadcurrently on the work machine can thus be obtained, but this eighthcontrol allows precise control to be effected by correcting theabsorbing torque τ in Formula (7).

Formula (7) is a formula for determining the differential pressure ΔPLSso that the absorbing torque τ on the PQ curve is not exceeded.Therefore, when the operation lever operation amounts are low in whichhorsepower limitation due to the PQ curve is not encountered, the engineoutput and the pump load are matched at a torque no more than theabsorbing torque τ (maximum value) and thus a flow rate corresponding tothe lever stroke can be fed. It is therefore managed to correct thetorque τ in Formula (7) to τ' so that the absorbing torque will decreaseas the operation amount detected values S1 and S2 decrease.

With this eighth control, therefore, a calculation formula or the likefor determining the corrected value τ' so that the torque τ is decreasedas the aperture area sum total detected value A decreases is stored inadvance in the memory, not shown in the figures, in the controller 33.The corrected value τ' is thus calculated based on the contents of thememory and on the detected values S1 and S2 of the operation amountsensors 45 and 46. In this case, the aperture area sum total A may bedetermined from the sum total of the operation amounts S1 and S2, or itmay be taken as the larger of the operation amounts S1 and S2.

The corrected torque τ' is thus used to correct Formula (8) to Formula(10), whereby the corrected differential pressure ΔPLS can be found.

    ΔPLS=({εE·τ'/(k·Pp'·A)}2, ΔPLS.sub.max)                                       (10)

A control signal for obtaining this corrected differential pressure ΔPLSis thus output to the control valve 36, and operability in the microvelocity region is further improved.

As described above, according to this embodiment, the engine rotationalspeed, hydraulic pump discharge pressure, and operation valve operationamounts are each detected, the absorbing torque of the hydraulic pump isset, and the differential pressure ΔPLS is varied based on a specificrelationship that is established among these detected values, the settorque, and the differential pressure ΔPLS; it is thus possible torealize optimal lever operability suitable for the present workingconditions, and thus possible to remarkably improve the workingefficiency.

This embodiment, furthermore, has been explained assuming cases inwhich, as shown in FIG. 7(c), the hydraulic pump 2 is subjected toequivalent horsepower control; however, this embodiment can of coursealso be applied to cases in which the hydraulic pump 2 is subjected tofixed torque control, as long as it is control that allows the engineoutput torque and the hydraulic pump absorbing torque to be matched.

An example of differential pressure ΔPLS control is described below.

With the ninth, tenth, and eleventh controls, described below, thestructure depicted in FIG. 12 is used as the hydraulic circuit. Thestructure depicted in FIG. 12 differs from FIG. 1 in the followingpoints.

Specifically, a pipe line 48 that connects the cylinder chamber on thelarge diameter side of the regulator 12 with the differential pressurecontrol valve 37 communicates with a torque control valve 47 forcontrolling the absorbing torque of the hydraulic pump 2, and the swashangle of the swash plate 2a is controlled by means of the differentialpressure control valve 37 and the torque control valve 47.

One end of the torque control valve 47 is connected via a spring 47c toa push member 12b that pushes a piston 12a of the regulator 12, and thepump pressure Pp in a pipe line 14 is applied as the pilot pressure to apilot port 47b at the other end. An electronic solenoid 47a ispositioned on the same side as this pilot port 47b, and a control signalfrom the controller 33 is sent to this solenoid 47a.

This torque control valve 47 controls the swash angle of the swash plate2a so that the pump absorbing torque does not exceed the torque τdesignated by the controller 33. Specifically, a control signal thatdesignates the torque τ is output from the controller 33 to the solenoid47a of the torque control valve 47, and the swash plate position inputvia the push member 47c, i.e., the valve position, is moved so that thetorque τ designated by the pump capacity D and the discharge pressure Ppinput via the pilot port 47b are not exceeded, and the swash plate 2a isthereby controlled.

Meanwhile, the controller 33 carries out calculation processing in themanner described below based on the various input signals, outputs thecontrol signal obtained as a result to the solenoid 36a of the controlvalve 36 and to the electromagnetic solenoid 47a of the torque controlvalve 47, and thereby controls the swash angle of the swash plate 2a ofthe hydraulic pump 2, i.e., the discharge amount D (cc/rev) of thehydraulic pump 2, via the differential pressure control valve 37, thetorque control valve 47, and the regulator 12.

In this case, the controller 33 outputs a control signal that sets theabsorbing horsepower at a fixed value, as described below, to the torquecontrol valve 47. Specifically, a control signal is output to the torquecontrol valve 37 so that the absorbing horsepower of the hydraulic pump2 becomes a fixed horsepower corresponding to the input work mode (M1 .. . ), and the swash plate 2a of the hydraulic pump 2 is thus controlledvia the torque control valve 37. In this way, the matching point movesto the point of optimal efficiency for the present load conditions (seeF in FIG. 7(c)).

Meanwhile, the controller 33 outputs a control signal to the controlvalve 36 so that the differential pressure ΔPLS set in the mannerdescribed below is obtained. Specifically, the controller 33 controlsthe pump absorbing horsepower as well as the differential pressure;here, the control pressure Pc applied to the pilot port 37a of thecontrol valve 37 varies according to the control signal sent to thesolenoid 36a of the control valve 36, and the differential pressure ΔPLSis thus varied. With this embodiment, this differential pressure ΔPLS isvaried according to the various control manners, as described below, andthis improves the operability of the operation levers, not shown in thefigures, of the operation valves 7 and 8.

The variable control of the differential pressure ΔPLS used in thehydraulic circuit of FIG. 12 is described in detail below.

Ninth Control

With this ninth control, as with the sixth control, the rotational speedεE of the engine i and the discharge pressure Pp of the hydraulic pump2, i.e., the load pressure PLS of the work machine actuators 3 and 4 andthe operation amounts S1 and S2 of the operation valves 7 and 8, areeach detected, the absorbing torque τ of the hydraulic pump 2 is setaccording to equivalent horsepower control based on the detectedrotational speed εE and the target rotational speed εTH of the engine 1,and the differential pressure ΔPLS is varied according to these detectedvalues and the torque set value τ, whereby control limiting theabsorbing torque of the hydraulic pump 2 is effected, inconveniencessuch as engine failure are prevented, and good lever operability isobtained.

Generally, a relationship such as that given in Formula (7) below holds,as described above, among the differential pressure ΔPLS, the absorbingtorque τ of the hydraulic pump 2, the aperture area sum total A of theoperation valves 7 and 8, the discharge pressure Pp of the hydraulicpump 2, and the rotational speed εE of the engine 1.

    √(ΔPLS)=εE·τ/(k·Pp·A) (7)

The maximum value for the discharge quantity Q of the hydraulic pump 2is naturally determined when the operation valves 7 and 8 are operatedup to the maximum operation amounts at the maximum rotational speed ofthe engine 1. The differential pressure ΔPLS obtained through Formula(7) by first determining this discharge quantity Q maximum value and bytaking the corresponding maximum differential pressure as ΔPLS_(max)must not exceed the maximum differential pressure ΔPLS_(max).Ultimately, the differential pressure ΔPLS is determined by means ofFormula (8).

    ΔPLS=min({εE·τ/(k·Pp·A}2, ΔPLS.sub.max)                                       (8)

The εE, Pp, and A of the function εE τ/(k Pp A) in Formula (8) can beobtained from the detected values of the corresponding sensors, and τ isobtained by setting the absorbing torque τ of the hydraulic pump 2according to equivalent horsepower control based on the detectedrotational speed εE and the target rotational speed εTH of the engine 1.The aperture area sum total A may be obtained as the sum of the outputsS1 and S2 of the operation amount sensors 45 and 46, or may be obtainedas the larger of the outputs S1 and S2 of the operation amount sensors45 and 46.

As described above, the set horsepower varies depending on the work mode(M1 . . . ) (the equivalent horsepower curve shown in FIG. 7(c) isdifferent), and the absorbing torque τ set thereby thus also varies, sothat the function εE·τ/(k·Pp·A) on the right side of Formula (7) isprepared for each work mode (M1 . . . ) as a function in which theengine rotational speed εE . . . are variables, the functioncorresponding to the selected work mode (M1 . . . ) is selected, and thedifferential pressure ΔPLS is calculated based on this selectedfunction. Furthermore, the function εE·τ/(k·Pp·A) can also be preparedfor each drive state of the work machine actuators 3 and 4, according towhich work machine is being driven in which direction. Since theabsorbing torque to be set varies depending on the drive state, it isnecessary, for example, in the case of boom elevation, to set theabsorbing torque τ high because the load is high, and in the case ofbucket operation, the absorbing torque may be set low because the loadis low. Which work machine is being driven in which direction canmoreover be detected based on the outputs of the operation amountdetection sensors 45 and 46.

Since the maximum differential pressure ΔPLS_(max) also varies dependingon the drive state of the work machine actuators 7 and 8 and on theselected work mode (M1 . . . ), it can also be determined based onthese.

With this ninth control, therefore, the aforementioned function isselected based on the selected work mode (M1 . . . ), the type ofcurrently driven actuators 3 and 4 detected by the operation amountsensors 45 and 46, and the drive direction thereof, and the substitutionof the engine rotational speed εE . . . into this selected functionallows the differential pressure ΔPLS of Formula (7) to be determined.Meanwhile, the differential pressure maximum value ΔPLS_(max) isdetermined based on the selected work mode (M1 . . . ), the type ofcurrently driven actuators 3 and 4 detected by the operation amountsensors 45 and 46, and the drive direction thereof, the smallerdifferential pressure ΔPLS is determined by Formula (8), and a controlsignal is output to the control valve 36 so that this determineddifferential pressure ΔPLS is obtained. Meanwhile, a control signal forobtaining the set absorbing torque τ (εE, εTH) is output to the torquecontrol valve 47, and the swash plate 2a is controlled by the torquecontrol valve 47 so that the absorbing torque τ is not exceeded.

The operating characteristics (a) and (c) of FIG. 2 express therelationship between the operation amounts S1 and S2 and the workmachine actuator drive velocities v1 and v2 for cases in which the loadis low and cases in which the load is high, respectively, according tothis ninth control; since the differential pressure ΔPLS decreases asthe load Pp increases according to Formula (7), there is no shift fromthe characteristics (a) to the characteristics (b), in which there is aconsiderable dead band, even when the load Pp is high, but there is ashift towards the characteristics (c), in which there is a low gradient,so that the dead band is kept low, as is the case with characteristics(a), in which the load Pp is low, and good operability is maintained.Moreover, since the equivalent horsepower control of the absorbingtorque of the hydraulic pump 2 is effected at the same time, noinconveniences such as engine failure occur. The characteristics (b)shown by the dotted line in FIG. 2, furthermore, depict cases in whichcontrol is not effected based on Formula (7), from which it is seen thatthe dead band is enlarged and the operability is degraded due to atorque limitation encountered when the load is high.

Tenth Control

With the ninth control, the differential pressure ΔPLS is variedaccording to Formula (7), and good operability that is suitable for theload currently on the work machine can thus be obtained; with this tenthcontrol, however, the aim is to realize highly precise control bycorrecting the load Pp of Formula (7).

FIG. 11(c) shows the relationship between the pump discharge pressure Ppand the pump discharge quantity Q; since the PQ curve is generallyapproached as the sum total A of the aperture areas decreases, i.e., asthe operation amounts S1 and S2 of the operation levers decrease,variations in the actual pressure Pp result, as shown by G, invariations in the quantity Q, and thus in variations in the differentialpressure, which has an adverse effect on operability.

Ultimately, as shown in FIG. 11(a), this tenth control allows the pumpdischarge pressure detected value Pp to be corrected so that thedischarge pressure Pp' gradually increases, as shown by the dot-dash-dotline I and the dotted line H, as the aperture area sum total Adecreases. In FIG. 11(a), the solid line J depicts the relationshipbetween the detected value Pp and the corrected value Pp' when theaperture area sum total A is at its maximum value A_(max) ; when theaperture area sum total A is at its maximum, there is no degradation inoperability, and the detected value Pp is thus not corrected. When theaperture area sum total A is higher than the minimum value A_(min) andlower than the maximum value A_(max), correction is performed, asindicated by the dotted line H, and when the aperture area sum total Ais at the maximum value A_(max) then, as shown by the dot-dash-dot lineI, the corrected value is made larger than in the case of the dottedline H, so as to cope with the degradation of operability.

The reason that the corrected value decreases as the detected valueincreases is that, since the variation width of the quantity Q decreasesas the pump pressure Pp increases, as is clear from FIG. 11(c), thevariation in the differential pressure decreases, and not that muchcorrection is required.

The contents of FIG. 11(a) may be expressed, as FIG. 11(b), in which therelationship among the pump pressure detected value Pp, the aperturearea sum total A, and the corrected value Pp' is represented as athree-dimensional map K, and correction may be carried out according tothis three-dimensional map K.

Accordingly, with this tenth control, the contents of FIG. 11(a) or FIG.11(b) are stored in advance in the memory, not shown in the figures, inthe controller 33. The corresponding corrected value Pp' in FIG. 11(a)or FIG. 11(b) may thus be fetched based on the detected value Pp of thepump pressure sensor 44 and the detected values S1 and S2 of theoperation amount sensors 45 and 46. In this case, the aperture area sumtotal A may be determined from the sum total of the operation amounts S1and S2, or may be determined to be the larger of the operation amountsS1 and S2.

The thus obtained corrected value Pp' is then used to correct Formula(8) to Formula (9) below.

    ΔPLS=min({εE τ/(k Pp' A)}2, ΔPLS.sub.max) (9)

A control signal for obtaining this corrected differential pressure ΔPLSis then output to the control valve 36. As a result, operability in themicro velocity region of the operation levers is further improved.

Eleventh Control

With the ninth control, the differential pressure ΔPLS is variedaccording to Formula (7), and good operability that suits the loadcurrently on the work machine can thus be obtained, but this eleventhcontrol allows precise control to be effected by correcting theabsorbing torque τ in Formula (7).

Formula (7) is a formula for determining the differential pressure ΔPLSso that the absorbing torque τ on the PQ curve is not exceeded.Therefore, when the operation lever operation amounts are low in whichthe horsepower limitation due to the PQ curve is not encountered, theengine output and the pump load are matched at a torque no more than theabsorbing torque τ (maximum value) and thus a flow rate corresponding tothe lever stroke can be fed. It is therefore managed to correct thetorque τ in Formula (7) to τ' so that the absorbing torque will decreaseas the operation amount detected values S1 and S2 decrease.

With this eleventh control, therefore, a calculation formula or the likefor determining the corrected value τ' so that the torque τ is decreasedas the aperture area sum total detected value A decreases is stored inadvance in the memory, not shown in the figures, in the controller 33.The corrected value τ' is thus calculated based on the contents of thememory and on the detected values S1 and S2 of the operation amountsensors 45 and 46. In this case, the aperture area sum total A may bedetermined from the sum total of the operation amounts S1 and S2, or itmay be taken as the larger of the operation amounts S1 and S2.

The corrected torque τ' is thus used to correct Formula (8) to Formula(10), whereby the corrected differential pressure ΔPLS can be found.

    ΔPLS=({εE·τ'/(k·Pp'·A)}2, ΔPLS.sub.max)                                       (10)

A control signal for obtaining this corrected differential pressure ΔPLSis thus output to the control valve 36, and operability in the microvelocity region is further improved.

As described above, according to this embodiment, the engine rotationalspeed, hydraulic pump discharge pressure, and operation valve operationamounts are each detected, the absorbing torque of the hydraulic pump isset, and the differential pressure ΔPLS is varied based on a specificrelationship that is established among these detected values, the settorque, and the differential pressure ΔPLS; it is thus possible torealize optimal lever operability suitable for the present workingconditions, and thus possible to remarkably improve the workingefficiency.

This embodiment, furthermore, has been explained assuming cases inwhich, as shown in FIG. 7(c), the hydraulic pump 2 is subjected toequivalent horsepower control; however, this embodiment can of coursealso be applied to cases in which the hydraulic pump 2 is subjected tofixed torque control, as long as it is control that allows the engineoutput torque and the hydraulic pump absorbing torque to be matched.

INDUSTRIAL APPLICABILITY

As described above, according to this invention, the differentialpressure is varied according to, among other things, the load on thework machine actuators, an optimal lever operability suitable for thecurrent working conditions can be realized, and a working efficiency isdrastically improved compared to that in the past.

According to this invention, moreover, the differential pressure isvaried so that the differential pressure decreases as the load on thework machine actuators increases, and as the engine rotational speeddecreases, and this makes it possible to avoid the effects of pressureoil leakage and to maintain good operability. According to thisinvention, moreover, the differential pressure is varied so that, whenthe operation valve is in a neutral position, the differential pressuredecreases than when the operation valve is operated, and so that thedifferential pressure decreases as the engine rotational speedincreases; this improves the operability when operation lever operationis begun, and also improves the working efficiency.

According to this invention, moreover, the differential pressure iscontrolled while the engine output torque and the hydraulic pumpabsorbing torque are matched; this eliminates inconveniences such asengine failure, and simultaneously improves operability.

According to this invention, moreover, since the differential pressureis controlled while the engine output torque and the hydraulic pumpabsorbing torque are matched, inconveniences such as engine failure areeliminated and, at the same time, the operability is improved.

We claim:
 1. A controller for a hydraulic drive machine, which has ahydraulic pump driven by a motor, a plurality of hydraulic actuatorsdriven by the supply of discharge pressure oil of the hydraulic pump viaa pressure oil supply line, and a plurality of flow rate control valvesfor controlling, in accordance with operation amounts, a flow rate ofthe pressure oil supplied to the plurality of work machine actuators;and which controls a discharge flow rate of the hydraulic pump so that adifferential pressure between a discharge pressure of the hydraulic pumpand a load pressure of the plurality of work machine actuators becomes aset value, characterised in that the controller comprises:pressuredetection means for detecting the discharge pressure of the hydraulicpump or the load pressure of the plurality of work machine actuatorsthroughout a pressure operating range; and means for varying thedifferential pressure set value substantially throughout the pressureoperating range so that the differential pressure set value decreases asthe pressure detected by the pressure detection means increases.
 2. Acontroller for a hydraulic drive machine, which has a hydraulic pumpdriven by a motor, a plurality of hydraulic actuators driven by thesupply of discharge pressure oil of the hydraulic pump via a pressureoil supply line, and a plurality of flow rate control valves forcontrolling, in accordance with operation amounts, a flow rate of thepressure oil supplied to the plurality of work machine actuators; andwhich controls a discharge flow rate of the hydraulic pump so that adifferential pressure between a discharge pressure of the hydraulic pumpand a load pressure of the plurality of work machine actuators becomes aset value, characterised in that the controller comprises:pressuredetection means for detecting the discharge pressure of the hydraulicpump or the load pressure of the plurality of work machine actuators;operation amount detection means for detecting the operation amounts ofthe plurality of flow rate control valves; and means for varying thedifferential pressure set value so that the differential pressure setvalue decreases as the pressure detected by the pressure detection meansincreases, and for varying the differential pressure set value so thatthe differential pressure set value decreases as the operation amountsdetected by the operation amount detection means increase, while thedifferential pressure set value determined by the pressure detectionmeans is being taken as a minimum value.
 3. A controller for a hydraulicdrive machine, which has a hydraulic pump driven by a motor, a pluralityof hydraulic actuators driven by the supply of a discharge pressure oilof the hydraulic pump via a pressure oil supply line, and a plurality offlow rate control valves for controlling, in accordance with operationamounts, a flow rate of the pressure oil supplied to the plurality ofwork machine actuators; and which controls a discharge flow rate of thehydraulic pump so that a differential pressure between a dischargepressure of the hydraulic pump and a load pressure of the plurality ofwork machine actuators becomes a set value, characterised in that thecontroller comprises:pressure detection means for detecting thedischarge pressure of the hydraulic pump or the load pressure of theplurality of work machine actuators; operation amount detection meansfor detecting the operation amounts of the plurality of flow ratecontrol valves; work machine type detection means for detecting the typeof work machine actuator currently being driven from among the pluralityof work machine actuators; and means for varying the differentialpressure set value based on the type of work machine actuator detectedby the work machine type detection means, on the operation amountsdetected by the operation amount detection means, and on the pressuredetected by the pressure detection means.
 4. The controller as definedin claim 3, characterised in that the work machine detection meansdetects the type of work machine actuator by detecting operatingconditions of the plurality of flow rate control valves.
 5. A controllerfor a hydraulic drive machine, which has a hydraulic pump driven by amotor, a plurality of hydraulic actuators driven by the supply of adischarge pressure oil of the hydraulic pump via a pressure oil supplyline, and a plurality of flow rate control valves for controlling, inaccordance with operation amounts, a flow rate of the pressure oilsupplied to the plurality of work machine actuators; and which controlsa discharge flow rate of the hydraulic pump so that a differentialpressure between the discharge pressure of the hydraulic pump and a loadpressure of the plurality of work machine actuators becomes a set value,characterised in that the controller comprises:pressure detection meansfor detecting the discharge pressure of the hydraulic pump or the loadpressure of the plurality of work machine actuators; rotational speeddetection means for detecting a rotational speed of the motor; and meansfor varying the differential pressure set value so that the differentialpressure set value increases as the rotational speed detected byrotational speed detection decreases and the pressure detected by thepressure detection means increases.
 6. A controller for a hydraulicdrive machine, which has a hydraulic pump driven by a motor, a pluralityof hydraulic actuators driven by the supply of a discharge pressure oilof the hydraulic pump via a pressure oil supply line, and a plurality offlow rate control valves for controlling, in accordance with operationamounts, a flow rate of the pressure oil supplied to the plurality ofwork machine actuators; and which controls a discharge flow rate of thehydraulic pump so that the differential pressure between a dischargepressure of the hydraulic pump and a load pressure of the plurality ofwork machine actuators becomes a set value, characterised in that thecontroller comprises:neutral position detection means for detecting afact that operating positions of the plurality of flow rate controlvalves have reached neutral positions; rotational speed detection meansfor detecting the rotational speed of the motor; and means for varyingthe differential pressure set value, when the operating positions of allof the plurality of flow rate control valves have been detected as beingin neutral positions by the neutral position detection means, so thatthe differential pressure set value is less than a differential pressureset value being set for a case when any of the plurality of flow ratecontrol valves is being operated, and so that the differential pressureset value decreases as the rotational speed detected by the rotationalspeed detection means increases.
 7. A controller for a hydraulic drivemachine, which has a hydraulic pump driven by a motor, a plurality ofhydraulic actuators driven by the supply of a discharge pressure oil ofthe hydraulic pump via a pressure oil supply line, and a plurality offlow rate control valves for controlling, in accordance with operationamounts, a flow rate of the pressure oil supplied to the plurality ofwork machine actuators; and which controls a discharge flow rate of thehydraulic pump so that a differential pressure between a dischargepressure of the hydraulic pump and a load pressure of the plurality ofwork machine actuators becomes a set value, characterised in that:therotational speed of the motor and either the discharge pressure of thehydraulic pump or the load pressure of the work machine actuators andthe operation amounts of the plurality of flow rate control valves aredetected, respectively, an absorbing torque of the hydraulic pump is setbased on a target rotational speed of the motor, and the differentialpressure is varied in accordance with each of the detected values andthe set value.
 8. The controller as defined in claim 7, characterised inthat the pressure detected value is corrected so that the load pressureof the plurality of work machine actuators or the discharge pressure ofthe hydraulic pump increases as the detected values of the operationamounts of the plurality of flow rate control valves decrease, and thedifferential pressure set value is varied according to the correctedpressure.
 9. The controller as defined in claim 7, characterised in thatthe absorbing torque set value is corrected so that the absorbing torqueof the hydraulic pump decreases as the detected values of the operationamounts of the plurality of flow rate control valves decrease, and thedifferential pressure set value is varied according to the correctedabsorbing torque.
 10. A controller for a hydraulic drive machine, whichhas a hydraulic pump driven by a motor, a plurality of hydraulicactuators driven by the supply of a discharge pressure oil of thehydraulic pump via a pressure oil supply line, and a plurality of flowrate control valves for controlling, in accordance with operationamounts, a flow rate of the pressure oil supplied to the plurality ofwork machine actuators; and which controls a discharge flow rate of thehydraulic pump so that a differential pressure between a dischargepressure of the hydraulic pump and a load pressure of the plurality ofwork machine actuators becomes a set value, characterised in that thecontroller comprises:work type designation means for selecting anddesignating the type of work performed by the hydraulic drive machine;rotational speed detection means for detecting the rotational speed ofthe drive machine; pressure detection means for detecting the loadpressure of the plurality of work machine actuators or the dischargepressure of the hydraulic pump; operation amount detection means fordetecting the operation amounts of the plurality of flow rate controlvalves; work machine type detection means for detecting the type of workmachine actuator currently being driven from among the plurality of workmachine actuators; torque setting means for setting a absorbing torqueof the hydraulic pump based on the type of work machine actuatordetected by the work machine type detection means, the type of workdesignated by the work type designation means, and a target rotationalspeed of the motor; and characterised in that the differential pressureset value is varied according to a torque set value set by the torquesetting means and each of the detected values detected by the respectivedetection means.
 11. A controller for a hydraulic drive machine, whichhas a hydraulic pump driven by a motor, a plurality of hydraulicactuators driven by the supply of a discharge pressure oil of thehydraulic pump via a pressure oil supply line, and a plurality of flowrate control valves for controlling, in accordance with operationamounts, a flow rate of the pressure oil supplied to the plurality ofwork machine actuators; and which controls a discharge flow rate of thehydraulic pump so that a differential pressure between a dischargepressure of the hydraulic pump and a load pressure of the plurality ofwork machine actuators becomes a set value, characterised in that thecontroller comprises:rotational speed detection means for detecting arotational speed of the motor; discharge pressure detection means fordetecting the discharge pressure of the hydraulic pump; load pressuredetection means for detecting the load pressure of the plurality of workmachine actuators; operation amount detection means for detecting theoperation amounts of the plurality of flow rate control valves; acontroller for setting an absorbing torque of the hydraulic pump basedon a target rotational speed of the motor, for setting the differentialpressure based on the operation amount detected values of the operationamount detection means, on the pressure detected value of the loadpressure detection means or the discharge pressure detection means, onthe rotational speed detected value of the rotational speed detectionmeans, and on the set absorbing torque, and for outputting controlsignals corresponding to the absorbing torque set value and thedifferential pressure set value; a torque control valve for controllinga swash angle of a swash plate of the hydraulic pump, so that theabsorbing torque set value is obtained, based on inputs from thecontroller of a control signal corresponding to the absorbing torque setvalue and a detection signal corresponding to the discharge pressuredetected value of the discharge pressure detection means; and adifferential pressure control valve for controlling the swash angle ofthe swash plate of the hydraulic pump, so that the differential pressureset value is obtained, based on the input from the controller of acontrol signal corresponding to the differential pressure set value, andon the input of detected signals corresponding to the pressure detectedvalues of the load pressure detection means and the discharge pressuredetection means.
 12. The controller as defined in claim 11,characterised in that the controller corrects the pressure detectedvalue of the discharge pressure or the load pressure so that thedischarge pressure or the load pressure increases as the detected valuesof the operation amounts of the plurality of flow rate control valvesdecrease, and sets the differential pressure according to the correctedpressure.
 13. The controller as defined in claim 11, characterised inthat the controller corrects the absorbing torque set value so that theabsorbing torque of the hydraulic pump decreases as the detected valuesof the operation amounts of the plurality of flow rate control valvesdecrease, and sets the differential pressure according to the correctedabsorbing torque.
 14. A controller for a hydraulic drive machine, whichhas a hydraulic pump driven by a motor, a plurality of hydraulicactuators driven by the supply of a discharge pressure oil of thehydraulic pump via a pressure oil supply line, and a plurality of flowrate control valves for controlling, in accordance with operationamounts, a flow rate of the pressure oil supplied to the plurality ofwork machine actuators; and which controls a discharge flow rate of thehydraulic pump so that a differential pressure between a dischargepressure of the hydraulic pump and a load pressure of the plurality ofwork machine actuators becomes a set value, characterised in that thecontroller comprises:work type designation means for selecting anddesignating the type of work performed by the hydraulic drive machine;rotational speed detection means for detecting a rotational speed of themotor; discharge pressure detection means for detecting the dischargepressure of the hydraulic pump; load pressure detection means fordetecting the load pressure of the plurality of work machine actuators;operation amount detection means for detecting the operation amounts ofthe plurality of flow rate control valves; work machine type detectionmeans for detecting the type of work machine actuator currently beingdriven from among the plurality of work machine actuators; a controllerfor setting an absorbing torque of the hydraulic pump based on the typeof work machine actuator detected by the work machine type detectionmeans, the type of work designated by the work type designation means,and a target rotational speed of the motor; for setting the differentialpressure based on the operation amount detected values of the operationamount detection means, the pressure detected values of the loadpressure detection means or the discharge pressure detection means, andthe set absorbing torque; and for outputting control signalscorresponding to the absorbing torque set value and the differentialpressure set value; a torque control valve for controlling a swash angleof a swash plate of the hydraulic pump, so that the absorbing torque setvalue is obtained, based on an input from the controller of a controlsignal corresponding to the absorbing torque set value, and on an inputof a detected signal corresponding to the discharge pressure detectedvalue of the discharge pressure detection means; and a differentialpressure control valve for controlling the swash angle of the swashplate of the hydraulic pump, so that the differential pressure set valueis obtained, based on the input from the controller of a control signalcorresponding to the differential pressure set value, and on the inputof detected signals corresponding to the pressure detected values of theload pressure detection means and the discharge pressure detectionmeans.